Transmissions for opposed-piston engines with two crankshafts

ABSTRACT

A transmission for an opposed-piston engine with two crankshafts includes a crankshaft gear train that combines the torque inputs from the two crankshafts and a gear arrangement coupled to the gear train that is operable to obtain various speed ratios for an output torque drive.

RELATED APPLICATIONS

This application is a divisional of U.S. patent application Ser. No.15/142,261, filed Apr. 29, 2016, titled “Transmissions ForOpposed-Piston Engines With Two Crankshafts,” published as U.S.Application Publication no. 2017-0314646 on Nov. 2, 2017. Thisapplication contains subject matter related to the subject matter of thefollowing commonly-owned U.S. patent applications: U.S. patentapplication Ser. No. 13/385,539, filed Feb. 23, 2012 for “DualCrankshaft, Opposed-Piston Engine Constructions,” now U.S. Pat. No.10,060,345; U.S. patent application Ser. No. 13/891,466, filed May 10,2013 for “Placement Of An Opposed-Piston Engine In A Heavy-Duty Truck,”now U.S. Pat. No. 9,849,770; U.S. patent application Ser. No.13/944,787, filed Jul. 17, 2013 for “Gear Noise Reduction inOpposed-Piston Engines,” U.S. Pat. No. 9,618,108; U.S. patentapplication Ser. No. 14/074,618, filed Nov. 7, 2013 for “Gear NoiseReduction in Opposed-Piston Engines,” published as U.S. ApplicationPublication no. 2015-0020629; and U.S. patent application Ser. No.14/450,747, filed Aug. 4, 2014 for “Split Gear Assembly With One-WayRoller Clutch For Controlling Backlash In Opposed-Piston Engines,” nowU.S. Pat. No. 9,772,030.

FIELD

The field relates to a transmission for an opposed-piston engine withtwo crankshafts which integrates the elements of a crankshaft gear trainwith elements of a multispeed transmission. More specifically, the fieldrelates to an opposed-piston engine equipped with a transmission gearbox that receives torque inputs from two crankshafts of the engine andprovides a multi-speed torque output drive.

BACKGROUND

FIG. 1 illustrates an arrangement of cylinders, pistons, and crankshaftsin an opposed-piston engine. The figure shows a three-cylinderarrangement, although this is not intended to be limiting; in fact, thebasic architecture portrayed in FIG. 1 is applicable to opposed-pistonengines with fewer, or more, cylinders. The opposed-piston engine 10includes cylinders 12, each including exhaust and intake ports 14 and16. Preferably, the cylinders comprise liners (also called “sleeves”)that are fixedly mounted in tunnels formed in an engine frame or block18. A pair of pistons (unseen in this figure) is disposed for opposingreciprocal movement in the bore of each cylinder 12. The opposed-pistonengine 10 includes an interlinked crankshaft system including tworotatably-mounted crankshafts 21 and 22 and a crankshaft gear train 30linking the crankshafts and coupling them to a power take-off shaft(“PTO shaft”). The crankshafts 21 and 22 are mounted to the engine bymain bearing arrangements (not shown), one at the bottom of the engineblock 18 and the other at the top. The crankshaft gear train 30 issupported in one end of the engine block 18 and is contained in acompartment 31 therein that is accessed through a removable cover 32.

As per FIG. 1, one piston of each piston pair is coupled to a respectivecrank journal 23 of the crankshaft 21 by a connecting rod assembly 27;the other piston is coupled to a respective crank journal 25 of thecrankshaft 22 by a connecting rod assembly 29. The crankshafts 21 and 22are disposed with their longitudinal axes in a spaced-apart, parallelarrangement. The crankshaft gear train 30 includes a plurality of gears,including two input gears 36 a and 36 b, which are fixed to respectiveends of the crankshafts 21 and 22 for rotation therewith. An output gearis mounted for rotation on a fixed shaft or post. The output gear 37drives a power take-off shaft 38 about an output axis of rotation A. Inthis configuration, two idler gears 39 a and 39 b are provided, eachmounted for rotation on a fixed shaft or post. The idler gear 39 ameshes with the input gear 36 a and the output gear 37; the idler gear39 b meshes with the input gear 36 b and the output gear 37. As a resultof the configuration of the crankshaft gear train 30, the crankshafts 21and 22 are co-rotating, that is to say, they rotate in the samedirection. However, this is not meant to so limit the scope of thisdisclosure. In fact, a gear train construction according to thisspecification may have fewer, or more, gears, and may havecounter-rotating crankshafts. Thus, although five gears are shown forthe crankshaft gear train 30, the numbers and types of gears for anyparticular crankshaft gear train are dictated only by the engine design.For example, the crankshaft gear train 30 may comprise one idler gearfor counter-rotation, or two idler gears (as shown) for co-rotation.

The gear train 30 shown in FIG. 1 represents a desirably convenient wayto connect two crankshafts of an opposed-piston engine for stableoperation and to unify the outputs of the crankshafts for delivery to adrive train via the power take-off shaft. In addition to a crankshaftgear train, a multispeed transmission is needed to convert the engine'soutput (speed and torque) as necessary to meet operating conditions of adrive train. In this regard, the term “transmission” also refers to adrive mechanism or gearbox comprising transmission gears arranged toselectably obtain speed ratios that match engine output to drive trainrequirements. In many instances, a gearbox matches engine output to thewheel speeds of a vehicle or a locomotive, or to propeller speeds of anaircraft or vessel. If the crankshaft gear train and the gearbox areprovided as separate units, with torque input to the gearbox via thepower take-off shaft, considerable redundancy is encountered inpackaging, and the length of the drive train is necessarily extended.Combining the crankshaft gear train and the gearbox into a single unitin which an arrangement of gears receives torque inputs directly fromthe spaced-apart crankshafts would offer potential benefits in reducingthe weight and size of the engine, and the length of the drive train.

U.S. Pat. No. 3,340,748 describes a multi-engine drive mechanism whichcouples two engines of an aircraft to drive a single propeller shaft byway of a single drive mechanism. The drive mechanism receives a torqueinput from one engine, or respective torque inputs from both engines,and allows the propeller shaft to be driven by either or both of theengines. The drive mechanism automatically establishes a predeterminedspeed ratio between the engines and the propeller shaft when bothengines operate at the same speed. When one engine ceases operation, thedrive mechanism automatically changes the speed ratio between the stilloperative engine and the propeller shaft in order to optimally drive theaircraft. Of course, the matter of driving a single output shaft fromtwo independent engines involves complex mechanical logic that must beable to combine torque inputs from independent sources and dynamicallyadapt to different torque input combinations. In any case, the speedratios remain fixed. In the case of a single opposed-piston engine withtwo crankshafts, however, the challenge is to combine twocontinuously-operating torque inputs in a single drive mechanismequipped to obtain various speed ratios.

SUMMARY

Thus, in some aspects, a transmission for an opposed-piston engine withtwo crankshafts (hereinafter, a “dual-crank opposed-piston engine”)includes a crankshaft gear train that combines the torque inputs fromthe two crankshafts and a gear arrangement coupled to the gear trainthat is operable to obtain various speed ratios with which to drive apower take-off shaft.

In other aspects, a transmission for a dual-crank opposed-piston enginecomprises a gear train for combining power from the crankshafts, a powertake-off shaft, and a gear arrangement interposed between the gear trainand the power take-off shaft for selectively providing output torque atone of a plurality of speed ratios. Preferably output torque is providedby means of a power take-off shaft.

In still further aspects, integrating the gear train of a dual-crankopposed-piston engine and with a transmission affords flexibility insupplying lubricating fluid for the engine and transmission. In somecases, the engine and transmission may be able to use the samelubricating fluid. In other cases, the transmission may require aseparate lubricating fluid, as when there may be a need to cool thetransmission lubricating fluid to a different extent than the enginelubricating fluid.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a side elevation view of a prior art arrangement of cylinders,pistons, and a gear train in a dual-crank opposed-piston engine.

FIG. 2A is a schematic representation of a prior art drive train for adual-crank opposed-piston engine. FIG. 2B is a schematic representationof a transmission for a dual-crank opposed-piston engine according tothis specification.

FIG. 3A is a schematic representation of a multi-speed parallel geararrangement integrated with the gear train of a dual-crankopposed-piston engine. FIG. 3B is a schematic representation of speedratios obtainable with the gear arrangement of FIG. 3A.

FIG. 4A is a schematic representation of a multi-speed planetary geararrangement integrated with the gear train of a dual-crankopposed-piston engine. FIG. 4B is a schematic representation of speedratios obtainable with the gear arrangement of FIG. 4A.

FIG. 5 is a schematic representation of the multi-speed parallel geararrangement of FIG. 3A with a reverse gear.

FIG. 6 is a schematic representation of the multi-speed gear arrangementof FIG. 4A with a reverse gear.

FIG. 7 is a schematic representation of a multi-speed hybrid geararrangement integrated with the gear train of a dual-crankopposed-piston engine.

FIG. 8 is a schematic representation of a multi-speed planetary geararrangement integrated with the gear train of a hybrid, dual-crankopposed-piston engine.

FIG. 9 is a schematic representation showing a common lubrication systemfor a transmission and a dual-crank opposed-piston engine,

FIG. 10 is a schematic representation showing separate lubricationsystems for a transmission and a dual-crank opposed-piston engine.

DETAILED DESCRIPTION

Referring to the prior art transmission arrangement for dual-crank,opposed-piston engine illustrated in FIG. 2A, the crankshaft 21 and thecrankshaft 22 of the engine 10 are individually rotated by movement ofopposed piston pairs 13 a/13 b in the cylinders 12. The crankshafts 21and 22 are coupled together, and their outputs are combined, by thecrankshaft gear train 30 which provides a single, unmediated torqueoutput drive via the power take-off (PTO) shaft 38. A multi-speedtransmission embodied in a gearbox 50 converts an output drive (speedand torque) presented on the PTO shaft 38 as necessary to meet operatingconditions of a drive train. A multi-speed torque output is provided tothe drive train on a transmission output shaft 51.

FIG. 2B schematically illustrates the union of the crankshaft gear train30 of the dual-crank opposed-piston engine 10 with a gear arrangement 60for a multi-speed transmission. In this union the transmission geararrangement 60 acts between the gear train 30 and the power take-offshaft 38, and is operable to provide a plurality of speed ratios withwhich to drive the power take-off shaft 38. This union eliminates theseparate transmission gearbox 50 and the transmission output shaft 51 ofthe prior art, and makes it possible to obtain different speeds forprovision to the drive train. The gear train compartment, now containingboth the gear train and a multi-speed gear arrangement, becomes atransmission gearbox 62.

The union of the crankshaft gear train and a multi-speed transmissiongear arrangement as per FIG. 2B may be achieved in multiple embodimentsby using parallel gears, planetary gears, or a combination of paralleland planetary gears. It enables many speed combinations, ranging fromminimum possible to maximum possible per packaging constraints, usingvarious gear combinations, In some embodiments, utilization of planetarygears also enables the coupling of the crankshaft gear train with anelectric motor, thereby supporting fuel-electric hybrid applications ofa dual-crank opposed-piston engine.

Example 1: FIG. 3A schematically illustrates a multi-speed parallel geararrangement integrated with the crankshaft gear train of a dual-crankopposed-piston engine in the transmission gearbox 62. In thisarrangement, the crankshaft gear train is disposed in parallel with agear arrangement for a multi-speed transmission. The crankshaft geartrain 30 is constructed as shown in FIG. 1 and a transmission geararrangement comprising a second gear train 70 is disposed such that thecrankshaft gear train 30 is located between the engine 10 and the secondgear train 70. The second gear train 70 comprises a set of transmissiongears including input gears 76 a and 76 b, output gear 77, and one ormore idler gears 79 a and 79 b. The input gear 76 a of the second geartrain is mounted to the end of the crankshaft 21, adjacent the inputgear 36 a of the crankshaft gear train, and the input gear 76 b of thesecond gear train is mounted to the end of the crankshaft 22, adjacentthe input gear 36 b of the crankshaft gear train. The output gear 77 ofthe second gear train, is rotatably mounted adjacent the output gear 37of the crankshaft gear train. The output gear 77 drives the rotatablePTO shaft 38. The second gear train 70 includes at least one idler gearfixed to the at least one idler gear of the crankshaft gear train 30 forrotation therewith. In this example, the second gear train 70 includes afirst idler gear 79 a fixed to the idler gear 39 a of the crankshaftgear train 30 for rotation therewith on a stationary post, and a secondidler gear 79 b fixed to the idler gear 39 b of the crankshaft geartrain 30 for rotation therewith on another stationary post.

The input gears 36 a and 76 a are conventionally mounted to the end ofthe crankshaft 21 by respective, separately actuated, frictionalcoupling mechanisms (not shown). These mechanisms may behydraulically-actuated and/or mechanically-actuated so as to permit eachinput gear to be locked onto and unlocked from the crankshaft,independently of the other input gear. When locked, an input gearrotates with the crankshaft; when unlocked, the input gear freewheels,driven by the idler with which it is meshed. The input gears 36 b and 76b are similarly mounted to the end of the crankshaft 22. The outputgears 37 and 77 are mounted to the PTO shaft 38 by respective,separately actuated, frictional coupling mechanisms (not shown). Whenlocked, an output gear rotates the PTO shaft 38; when unlocked, theoutput gear freewheels, driven by an idler with which it is meshed. Thefrictional coupling mechanisms are conventional, and are conventionallyactivated. For example, the frictional coupling mechanisms may comprisehydraulically-actuated cone clutches.

The gear sizes in the crankshaft gear train 30 and the second gear train70 are determined according to design requirements. In the example ofFIG. 3A, the input and output gear sizes of the crankshaft gear trainare larger than the input and output gear sizes of the second gear train70. The idler gears 39 a and 39 b of the crankshaft gear train 30 are ofa smaller gear size than the idler gears 79 a and 79 b of the secondgear train 70. The frictional coupling mechanisms that act between thecrankshafts and the input gears and between the output gears and the PTOshaft 38 are conventionally operated to obtain various speed ratiosbetween the input and output gears, and thus between the torque inputsof the two crankshafts and the torque output of the PTO shaft 38. Inthis regard, refer to FIG. 3B in which a horizontally shaded gear islocked to a shaft and an unshaded gear is unlocked and freewheels. Withreference to FIG. 3B and Table 1, a first speed (Speed A) is obtained bylocking the input gears 36 a and 36 b to the crankshafts 21 and 22(lock/lock), locking the output gear 37 to the PTO shaft 38 (lock),while unlocking the input gears 76 a and 76 b from the crankshafts 21and 22 (unlock/unlock) and unlocking the output gear 77 from the PTOshaft 38 (unlock). As per Table 1, four speeds are obtained with theparallel gear train arrangement of FIG. 3A.

TABLE 1 Input Output Input Gears 36a/36b Gear 37 Gears 76a/76b OutputGear 77 Speed A lock/lock lock unlock/unlock unlock Speed Bunlock/unlock lock lock/lock unlock Speed C unlock/unlock unlocklock/lock lock Speed D lock/lock unlock unlock/unlock lock

Example 2: FIG. 4A schematically illustrates a multi-speed planetarygear arrangement integrated with the gear train of a dual-crankopposed-piston engine. Multiple output speeds may be obtained byreplacing one or more gears of the crankshaft gear train with planetarygears. In FIG. 4A the crankshaft gear train 30 p comprises anarrangement of planetary gears separated by idler gears. In this examplethree gears—the input gears 36 a and 36 b, and the output gear 37—arereplaced by two-speed planetary input gears 36 ap, 36 bp, and aplanetary output gear 37 p, to obtain four output speeds. Preferably,each of the planetary input gears 36 ap and 36 bp includes a planetcarrier (CAR) fixed to the end of a respective crankshaft (CAR to CRNK),a ring gear (RNG) meshed on its outer annular face with an adjacentidler gear and on its inner annular surface with the planetary gears ofthe planet carrier, and a sun gear (SUN) held stationary on a fixed post(GND). Each planetary input gear includes a hydraulically-actuated or amechanically-actuated frictional coupling mechanism (not shown) thatpermits the ring gear and planet carrier to be locked together so as torotate at the same speed and unlocked so as to rotate at differentspeeds. Thus, each planetary input gear 36 a and 36 b converts the inputspeed of the crankshaft to which its planet carrier is fixed to a firstoutput speed of its ring gear when its ring gear and planet carrier arelocked together and to a second output speed of its ring gear when itsring gear and planet carrier are unlocked. In the example shown in FIG.4B, the speed ratio thus obtained by each planetary input gear is 1:1when the ring gear and planet carrier are locked; when the ring gear andplanet carrier are unlocked, the speed ratio is 1:1.22. The speed ratiothus obtained by the planetary output gear is 1:1 when the ring gear andplanet carrier are locked; when the ring gear and planet carrier areunlocked, the speed ratio is 1:1.5.

With reference to FIG. 4A, the planetary output gear 37 p includes aplanet carrier (CAR) fixed to the PTO shaft 38 (CAR to PTO), a ring gear(RNG) meshed on its outer annular face with adjacent idler gears 39 aand 39 b (RNG to IDLER), and on its inner annular surface with theplanetary gears of the planet carrier, and a sun gear (SUN) heldstationary on a fixed post (GND). The planetary output gear 37 pincludes a hydraulically-actuated or a mechanically-actuated frictionalcoupling mechanism (not shown) that permits the ring gear and planetcarrier to be locked together so as to rotate at the same speed andunlocked so as to rotate at different speeds. In the example shown inFIG. 4B, x is an input speed at which the crankshafts rotate. The PTOshaft 38 rotates at an output speed of x (the input speed) multiplied byfactor having a value determined a combination of locked and unlockedconditions of the planetary gears.

As per FIGS. 4A and 4B, the frictional coupling mechanisms that actbetween the planet carriers and ring gears of the planetary input gears36ap and 36bp and between the planet carrier and ring gear of theplanetary output gear 37p are conventionally operated in order to obtainvarious speed ratios between the planetary input and output gears, andthus between the torque inputs of the two crankshafts and the torqueoutput of the PTO shaft 38. In this regard, with reference to FIG. 4Band Table 2 a first speed (Speed A) is obtained by locking the planetcarrier to the ring gear (lock CAR/RNG) of both planetary input gears 36ap and 36 bp, and locking the planet carrier to the ring gear (lockCAR/RNG) of the planetary output gear 37 p. As per FIG. 4B and Table 2,four speeds are obtained with the planetary arrangement of FIG. 4A.

TABLE 2 Input Gear 36ap Output Gear 37 Input Gear 36bp Speed A lockCAR/RNG lock CAR/RNG lock CAR/RNG Speed B lock CAR/RNG unlock CAR/RNGlock CAR/RNG Speed C unlock CAR/ lock CAR/RNG unlock CAR/RNG RNG Speed Dunlock CAR/ unlock CAR/RNG unlock CAR/RNG RNG

Example 3: As seen in FIG. 5, the parallel gear train arrangement ofFIG. 3A may be equipped with a reverse gear arrangement by mounting agear 90 fixed to rotate with a pair of idler gears on a stationary post.The parallel gear train arrangement of FIG. 5 has a first stateproviding a forward gear range output and a second state providing areverse gear output. A smaller, reverse gear 92 meshed with the gear 90is mounted to rotate on a stationary post. The output gear 77 of thesecond gear train 70 is mounted on a moveable spline shaft 93. Ahydraulically actuated and/or mechanically-actuated means is provided tomove the spline shaft 93 into and out of engagement with the reversegear 92 as shown by the arrow 94. When the spline shaft 93 disengagesthe reverse gear 92, the parallel gear arrangement provides a forwardgear range that comprises the four speeds illustrated in FIG. 3B andTable 1. When the spline shaft 93 engages the reverse gear 92, theparallel gear arrangement provides a reverse gear output on the PTOshaft 38.

Example 4: As seen in FIG. 6, the gear train of FIG. 4A may be modifiedwith a reverse gear arrangement in which a fourth planetary gear 95 isdisposed concentrically with the planetary gear 37 p in adouble-planetary gear drive configuration that constitutes the output ofthe modified crankshaft gear train. The double-planetary driveconfiguration has a first state providing a forward gear range output,and a second state providing a reverse gear output. The double-planetarygear drive configuration has two stages. The first stage comprises theplanetary gear 37 p, and the second stage comprises the planetary gear95. The two-stage arrangement includes first and secondhydraulically-actuated or mechanically-actuated frictional couplingmechanisms 96 and 97. The frictional coupling mechanism 96 (for example,a clutch) permits the grounded sun gear of the first stage 37p to belocked to and unlocked from the planet carrier of the second stage 95.The frictional coupling mechanism 97 (for example, a clutch) permits thering gear of the second stage 95 to be locked to and unlocked from theplanet carrier of the first stage 37 p. The planet carrier of firststage 37 p is rigidly connected to the sun gear of second stage 95. Theoutput of the gear-train is taken from the ring gear in the second stage95, which is coupled to the PTO shaft 38 (RNG to PTO 38). For theforward gear range output, the ring gear of the second stage 95 islocked by 97 to the planet carrier of the first stage 37 p. The planetcarrier of the second stage 95 is unlocked by 96 from the grounded sungear of the first stage 37 p, whereby the whole second stage 95 rotatesas a single unit at one of the four forward speeds determined by a geartrain setting illustrated in FIG. 4B. When the planet carrier of thesecond stage 95 is held stationary by being locked to the sun gear ofthe first stage 37 p, the ring gear of the second stage 95 reversesdirection and provides reverse gear output on the PTO shaft 38 thatrotates at Z_(SUN)/Z_(RNG) times the speed of the sun gear, whereZ_(SUN) is the tooth count of the sun gear of the second stage 95 andZ_(RNG) is the tooth count of the ring gear of the second stage 95.

Example 5: As seen in FIG. 7, a two-speed planetary gear 100 may beadded to the parallel gear train arrangement of FIG. 3A to increase thenumber of speed ratios available from the gearbox. In thisconfiguration, the sun gear of the planetary gear 100 is rigidlyconnected to the shaft which is driven by one or the other of the outputgears 37 and 77. Thus, the sun gear is driven by the crankshafts (SUN toCRNK) and rotates at the four speeds illustrated in FIG. 3B and Table I.The ring gear of the two-speed planetary gear 100 is grounded (RNG toGND), and so does not rotate. The planet carrier of the two-speedplanetary gear 100 is rigidly connected to the PTO shaft 38 (CAR toPTO). The two-speed planetary gear 100 includes a hydraulically-actuatedor a mechanically-actuated frictional coupling mechanism (not shown)that permits the sun gear and planet carrier to be locked together so asto rotate at the same speed and unlocked so as to rotate at differentspeeds. When the sun and planet carrier are locked together, the PTOshaft 38 rotates at the four speeds illustrated in FIG. 3B and Table I.When the sun and planet carrier are unlocked, the PTO shaft 38 rotatesat four additional speeds related to the four speeds illustrated in FIG.3B and Table I by Z_(SUN)/Z_(PLANET) times the speed of the sun gear,where Z_(SUN) is the tooth count of the sun gear and Z_(PLANET) is thetooth count of a planet gear carried by the planet carrier.

Example 6: As seen in FIG. 8, the planetary gear arrangement of FIG. 4Aenables the coupling of the crankshaft gear train 30 p with an electricmotor 110, thereby supporting fuel-electric hybrid applications of adual-crank, opposed-piston engine. The electric motor 110 is connectedto the sun gear of PTO shaft planetary output gear 37 p with a one-wayclutch (SUN to GND/ELECTRIC MOTOR). When the electric motor 110 does notrotate, the sun gear is grounded (stationary). When the motorover-rotates ground, which is to say that the angular velocity (in RPM)of the motor is greater than zero, speed is added to the PTO shaft 38 ina continuously variable way. The speed of the electric motor 110 isindependent of the engine 10 speed. Generally, the combined output speed(engine+motor) at the PTO shaft 38 is given by:

ω_(CAR)=(Z _(SUN)*ω_(SUN) +Z _(RNG)*ω_(RNG))/(Z _(SUN) +Z _(RNG)) . . .Where:

-   ω_(CAR): angular velocity of planet carrier-   Z_(SUN): tooth count of SUN gear-   ω_(SUN): angular velocity of Sun-   Z_(RNG): tooth count of RING gear-   ω_(RNG): angular velocity of RING gear-   Manifestly, when ω_(SUN)=0, the speed ratios of FIG. 4B and Table 2    are obtained.

Lubrication Options: Integrating crankshaft and transmission gear trainsprovides an opportunity for using a single fluid (engine oil) tolubricate both the engine and transmission. In cases where thetransmission requires a separate fluid, there may be a need to cool thetransmission fluid and the integrated crankshaft-transmission gear trainsystem has the potential to simplify the transmission fluid coolingcircuit.

With reference to FIG. 9, a common lubricating fluid (oil, for example)may be used to lubricate both the engine and the transmission. As shownin the circuit diagram, a single engine lubricating fluid can be sharedbetween engine and transmission by way of a lubrication systemcomprising an oil delivery network 130. The network 130 comprises amechanically-driven pump 132 that pressurizes oil collected in a commonsump 134 and delivers pressurized oil through a pressure relief valve136, a filter 138, and a cooler 140 to an engine oil header 142. Thepressure relief valve 136 regulates the oil pressure by returning oil tothe sump 134. The header 142 distributes the pressurized oil to variouselements of the engine 10, including the gearbox 62 that cabins theunified crankshaft/transmission gear assembly.

Synthetic, mineral, multi-grade or straight SAE grade engine oils arepotential candidates for use as common lubricating fluid for the engineand the transmission. The common lubricating fluid may provide severaladvantages such as a simpler lubricating circuit, common accessoriessuch as oil sump/filter/oil cooler/valves, fewer dynamic oil seals, lowfriction losses, ease of maintenance, one oil change interval, lowengine weight, and low cost. A disadvantage of using common lubricatingoil may be accelerated oil shearing due to gear mesh and roller bearingsloads. Oil shearing causes the oil to thin down to a grade that weakensthe oil film thickness between rubbing components and eventually leadsto inadequate lubricating properties for engine components. Special oilformulations can be used to provide high shear stability and also bettergear shift in the integrated engine/transmission application.

With reference to FIG. 10, separate lubricating fluids may be providedthrough separate oil supply networks 150 and 160. The oil supplynetworks 150 and 160 are essentially duplicates of the network 130 ofFIG. 9, with the exception that the oil supply network 150 supplies oneoil to lubricate the engine 10 and the oil supply network 160 suppliesanother, different oil to lubricate the gearbox 62. Dynamic seals 165are seated in those openings where the end sections of the crankshafts21 and 22 extend from the engine 10 into the gearbox 62. Each dynamicseal 165 acts between a respective one of the crankshafts 21 and 22 anda wall of the engine block to maintain separation of the two differentlubricating fluids in the engine.

Industrial Applications: In the examples and embodiments thus fardescribed, the dual-crank, opposed-piston construction is characterizedby parallel alignment of the crankshafts with the longitudinal axes ofthe cylinders extending orthogonally to, and intersecting, thelongitudinal axes of the crankshafts. In this construction each pistonis coupled to only one of the two crankshafts. The principles of gearboxconstruction according to this specification are not intended to belimited to this dual-crankshaft construction, and they may be appliedalso to other dual-crankshaft, opposed-piston engines. In someinstances, the cylinders and crankshafts may be arranged such that eachpiston is coupled to both crankshafts. One such example is described inU.S. Pat. No. 8,539,918 which is commonly-owned herewith.

Principles of gearbox construction for dual-crankshaft, opposed-pistonengines have been described with reference to certain embodiments andexamples. However, it should be understood that various modificationscan be made without departing from the scope of these principles.Accordingly, the scope of patent protection for these principles islimited only by the following claims.

What is claimed is:
 1. A transmission for an opposed-piston engine withtwo spaced-apart crankshafts, comprising: a gear train that links thetwo crankshafts together; a rotatable power take-off shaft; and, a geararrangement acting between the gear train and to the power take-offshaft that is operable to provide a plurality of speed ratios with whichto drive the power take-off shaft.
 2. The transmission of claim 1, inwhich the gear train comprises input gears, each input gear mounted toan end of a respective crankshaft, an output gear coupled to the powertake-off shaft, and at least one idler gear mounted for rotation betweenone of the input gears and the output gear.
 3. The transmission of claim2, in which the gear arrangement comprises one or more planetary gearsmeshed with the gear train.
 4. The transmission of claim 3, in whichplanetary gears mounted to the crankshafts comprise two-speed planetarygears.
 5. The transmission of claim 2, in which the gear arrangementcomprises one or more of the input gears being a planetary gear and theoutput gear being a planetary gear.
 6. The transmission of claim 5, inwhich the gear arrangement comprises both of the input gears beingtwo-speed planetary gears and the output gear being a two-speedplanetary gear.
 7. The transmission of claim 5, in which the geararrangement comprises a double-planetary gear drive having a first stateproviding a forward gear range output and a second state providing areverse gear output.
 8. The transmission of claim 2, in which the geararrangement comprises a set of transmission gears disposed in parallelwith the gear train and a planetary gear mounted to the power take-offshaft.
 9. The transmission of claim 8, in which the planetary gearcomprises a two-speed planetary gear.
 10. The transmission of claim 2,in which the gear arrangement comprises a reverse gear arrangement. 11.An opposed-piston engine with two crankshafts coupled to a transmissionaccording to any one of claims 1-4.
 12. A combination comprising anopposed-piston engine with two crankshafts coupled to the transmissionof claim 5, an electric motor and a frictional engagement devicecoupling the electric motor to a sun gear of the power take-offplanetary gear.